Pilot flow pressure control valves are extensively used for controlling the pressures at which a larger control valve will open to relieve or control the pressure in a hydraulic system. Normally, such pilot valves are comprised of a valve element having a surface or surfaces exposed to the hydraulic pressure to be controlled which creates a pressure force tending to open the valve. A compression spring biases the valve element to the closed position. When the pressure force exceeds the spring force or bias, the valve opens to allow the flow of fluid therethrough. This fluid flow then changes pressure forces on the larger valve allowing it to open and prevent further increases in the fluid pressure.
To determine the opening pressure, the amount of spring compression or bias force on the valve element is initially adjusted at the time of manufacture or is controlled in the field by means of a threaded member which is rotated either manually or by an electric motor.
The use of springs to determine the opening pressure of the valve presents three problems. First, manual adjustment prevents remote control and electric motor adjustment is expensive, bulky, relatively slow to react and requires relatively large amounts of pulsed electrical power. Computer control of such valves in industrial processes is difficult.
Secondly, springs inherently give to the valve a rising pressure versus flow characteristic. Thus, as the valve opens to permit flow, the spring is compressed. As the spring is compressed, its bias force increases. A progressively higher fluid force or fluid pressure is required to open the valve further and further. For example, with a spring having a force versus compression rate (hereinafter called "Spring rate") of 30 pounds per inch, to fully open a pressure control valve 0.020 inches, requires an increase in force of 0.60 pounds. With a 0.030 inch diameter valve seat, this calculates to a pressure increase of 84.8 pounds per square inch (psi) to fully open the valve. With a valve seat of 0.040 inches, the pressure increase calculates to 47.7 psi. This results in rising pressure versus flow characteristic which is considered undesirable for accurate pressure control. It is particularly a problem at the lower inlet pressures where 47.7 psi is about 47% of 100 psi.
Thirdly, unless an expensive and complicated compound wound spring is employed, it is almost impossible to provide a spring biased pilot relief valve which will perform equally well at both high and low pressures. Thus, at low pressures, e.g., 100 psi, the valve must open the maximum designed amount to permit the necessary fluid flow to effect operation of the main relief valve. At the higher pressures, e.g., 6,000 psi, the valve need only open a very small amount to permit the necessary volume of fluid flow to actuate the main relief valve. Thus, if a low spring rate spring is employed to give good pressure versus flow characteristics at low pressures, when the spring is compressed to require a higher opening pressure the valve becomes unstable.
If a high spring rate spring is employed to give stability and good pressure versus flow characteristics at high pressure, then the pressure versus flow characteristics at low pressure are poor.
At the higher pressures, this adverse effect of a rising force displacement characteristic of a spring, is exacerbated by the apparent drop in pressure on the valve element as the valve opens due to the conversion of the pressure energy on the element to velocity energy. This effect will hereinafter be referred to as the "Bernoulli Effect." The Bernoulli Effect results in an apparent drop of up to about 10% in the opening pressure force on the valve element as the flow increases. This apparent lowering of the pressure force against the increasing spring bias further adversely effects the pressure versus flow characteristics of the valve.
Ser. No. 378,133, filed May 14, 1982 attempted to overcome many of these problems. In that application, a pilot flow relief valve is described, comprised of: a housing defining an elongated cylindrical cavity having at one end an inlet port defined by a valve sea; a solenoid surrounding a portion of the housing defining the cavity; and a magnetically permeable armature slidable in the cavity, which when the solenoid is energized, biases a valve element having a valve seat engaging surface toward the valve seat with a magnetic force proportional to the degree of energization of the solenoid. When fluid pressure on the valve element creates a pressure force greater than the magnetic force, the valve seat engaging surface is forced away from the valve seat to allow fluid to flow through the inlet port into the cavity and out through a discharge port located adjacent to the inlet port, the amount of flow being proportional to the inlet pressure and the amount of movement of the valve seat engaging surface away from its valve seat.
In that valve, the housing inside of the solenoid is comprised of: a magnetically permeable armature attracting sleeve adjacent the valve seat; a magnetically permeable armature supporting sleeve spaced from the attracting sleeve and remote from the valve seat; and, an intermediate magnetically nonpermeable sleeve between the two magnetically permeable sleeves forming a magnetic airgap therebetween.
The armature has a supported end remote from the valve seat slidable in close spaced relationship with and in substantial overlapping relationship with the supporting sleeve and an attracted end adjacent the valve seat slightly overlapping the end of the attracting sleeve adjacent the airgap by a predetermined amount which end is externally tapered toward the valve seat. In addition, the end of the supporting sleeve remove from the valve seat also includes an axially facing, magnetically permeable surface axially spaced a controlled distance from the corresponding end of the armature.
The two magnetically permeable sleeves and the axially facing surface form a shaped magnetic circuit for the solenoid, which circuit, for any fixed degree of solenoid energization, provides a magnetic force on the armature toward the valve seat which, for any fixed solenoid energization, increases as the opening of the valve forces the armature away from the valve seat and decreases its overlap with its attracting sleeve.
A plot of the magnetic force versus the displacement of the attracted end of the armature relative to the attracting end of the attracting sleeve, provides a force displacement curve which as the ends are moved from a negative overlap (axially spaced) to a zero overlap (when the ends are exactly aligned) to full overlap, first rises rapidly as the overlap approaches zero, flattens in a broad curve to a maximum at approximately zero overlap and then decreases on a generally linear curve to zero as the overlap increases to the maximum.
For proper operation of the valve at all pressures, the initial (or maximum) overlap of the two ends must be adjusted such that: at all solenoid energizations (within the designed range of energization), the segment of the curve of a length corresponding to the designed maximum movement of the valve element away from the valve seat, is always rising. There must always be an overlap.
A segment of the force-displacement curve is selected of a length corresponding to the maximum designed opening movement of the valve which provides the maximum possible bias force while still increasing linearly as the valve element is forced away from the valve seat by increasing hydraulic pressures.
Further testing indicated that as the solenoid energizing voltage increased, the point in the overlap where the maximum force occurs varies, with the maximum forces at higher energizations occurring at the greater degrees of overlap. This means that at the higher relief pressures, (i.e. a higher solenoid energization) a different initial overlap is required than at the lower relief pressures i.e., lower solenoid energization. Such initial overlaps must be fixedly adjusted at the time of manufacture. It was found that if the overlap was adjusted to give the desired slope for the low pressures, the curve sloped downwardly at high pressures. The valve was thus unstable at high pressures. If the initial overlap was adjusted to give a desired slope for high pressure, then at the lower degrees of solenoid energization there was insufficient force for satisfactory operation at the lower pressures.
It was determined that it was necessary to provide, at all degrees of solenoid energization, a magnetic force displacement curve which slopes upwardly through the desired range of movement of the valve element, which in the case of the preferred embodiment is 0.020 inches.
While the stability problem could have been handled by adding a higher rate spring, acting in conjunction with the magnetic forces, this would have made even worse the problem of obtaining the desired volume of fluid flow at low pressure.